Axial fluid-film bearing

ABSTRACT

A hydrodynamic axial or thrust bearing for rotating spindles and the like comprising a cylindrical shell surrounding and spaced from the spindle. The spindle is provided with an annular flange fixed to it, which is located between two ring shaped discs supported by the shell. One of the discs is fixed against axial and rotational movement while the other is fixed against rotation but is supported to move axially. The movable disc is provided with resilient means normally biasing it toward the flange. Oil is supplied between the faces of the flange and discs.

United States Patent [19] Linda et al.

[ AXIAL FLUID-FILM BEARING [75] Inventors: Josef Linda; F rantisekRosberg;

Jaroslav Marsalek, all of Prague, Czechoslovakia [73] Assignee: TosHostivar narodni podnik,

Praha-Hostivar, Czechoslovakia [22] Filed: Apr. 7, 1971 [2i] Appl. No.:131,880

[30] Foreign Application Priority Data Apr. 13, 1970 Czechoslovakia245270 [52] US. Cl. 308/160 Int. Cl. Fl6c 17/04 [58] Field of Search308/l60, 9, 122

[ 56] References Cited UNITED STATES PATENTS 2,822,223 2/l958 Offen308/l68 [451 Sept. 25, 1973 1,325,459 12/1919 Wingfield 308/160 PrimaryExaminer-Charles J Myhre Assistant E.raminer-Frank SuskoAttorney-Richard Low and Murray Schaffer t [57] ABSTRACT A hydrodynamicaxial or thrust bearing for rotating spindles and the like comprising acylindrical shell surrounding and spaced from the spindle. The spindleis provided with an annular flange fixed to it, which is located betweentwo ring shaped discs supported by the shell. One of the discs is fixedagainst axial and rotational movement while the other is fixed againstrotation but is supported to move axially. The movable disc is providedwith resilient means normally biasing it toward the flange. Oil issupplied between the faces of the flange and discs.

12 Claims, 6 Drawing Figures PATENTEU SHEU 2 0F JosEF LINDA FfmwngekRossm JfiflosLm/ mtk lk INVENTOR 9 A ORNEY PATENTEU 3. 76 1.150

sum u m a INVENTORS ATT RNE AXIAL FLUID-FILM BEARING BACKGROUND OFINVENTION The present invention relates to spindle bearings and inparticular, to a hydrodynamic, axial fluid bearing for absorbing thrustloads.

Hydrodynamic thrust bearings have been used for sometime to journalrotating spindle shafts. Such bearings are very effective in journalinghigh speed shafts; providing accurate location of the shaft, lowvibration and extremely low frictional effect even under high loads.Greater attention has been given, however, to the construction of radialjournal bearings than to the precise construction of axial thrustbearings. This is a result of the fact that radial journal bearings aremore critical than the thrust bearings since the thrust load of aspindle or rotating shaft is not as great as the load radially. However,at the increased speed at which machine tools are now being operated,this problem has become more acute.

Axial hydrodynamic thrust bearings fall into two basic groups dependingon their operational function. The first group includes those structureswhich have fixed nonvariable clearances between the relatively rotatingparts, in which the fluid bearing film is formed, which clearance doesnot vary in response to changes in speed of spindle rotation. The secondgroup includes those in which the axial clearance between the same partsis adjusted in dependence on the speed of rotation of the spindle. Ineither case the bearing film is formed by compressing the fluid inentering wedges or by tilting or pivoting variously arranged bearingsegments.

The optimum clearance in bearings of the'first group must be achieved byvery precise manufacture or by means of complex and difficult adjustmentof the parts when assembled. The factors involved in obtaining suchoptimum arrangement are many and variable, namely; the slide velocity ofthe spindle and bearing, the inherent resiliency of the parts,bearing-capacity, heat generating and manufacturing tolerance. Mostimportant, however, is the skill and experience of the machine operatorand assembler of the bearings. If the bearings were adjusted for optimumoperation at normal speeds and if the speed of the spindle shouldsuddenly change or extend beyond the projected range, the bearing couldnot compensate nor could it prevent seizure or other damage to it. Onthe other hand if the bearings were set for high speeds it would notfunction efficiently at the lower and more normal speeds.

While the provision of the second group of bearings haslargelyeliminated the problems of the first group, they too havegenerated their own drawbacks. Operational elasticity (i.e.,adjustability) has been so far achieved by complicated and expensivestructural arrangement and the necessity of maintaining very hightechnical and manufacturing precision. These are serious drawbacks tothe wide use of such bearings.

It is an object of the present invention to provide a hydrodynamicthrust bearing which overcomes the drawbacks of the prior art.

It is another object of the present invention to pro-' vide ahydrodynamic thrust bearing which is automatically adjustable to varyingspeeds and spindle loads.

It is another object of the present invention to provide a hydrodynamicthrust bearing which is capable of exerting maximum rigidity (i.e.,elastic limits) under varying loads conditions.

. It is another object of the present invention to provide ahydrodynamic thrust hearing which is simple in design, which does notrequire high manufacturing tolerances and which is inexpensive.

These objects as well as others together with numerous advantages willbe apparent from the following disclosure of the present invention.

SUMMARY OF INVENTION According to the present invention a hydrodynamicaxial or thrust bearing for rotating spindles and the like is providedcomprising a cylindrical shell surrounding and spaced from the spindle.The spindle is provided with an annular flange fixed to it, which islocated between two ring shaped discs supported by the shell. One of thediscs is fixed against axial and rotational movement while the other isfixed. against rotation but is supported to move axially. The movabledisc is provided with resilient means normally biasing it toward theflange. Oil is supplied between the faces of the flange and discs. Whenthe spindle begins to rotate an oil film is formed between the faces andthe bearing assumes a balanced condition wherein the films are of thesame thickness. However, as the speed of the spindle increases or theload changes, the spindle and flange move axially, sliding the movabledisc. A new balanced condition occurs wherein the films between therespective faces are not equal in thickness.

Preferably the resilient member biasing the slidable disc is a springwasher of predetermined spring rate BRIEF DESCRIPTION OF DRAWINGS In theaccompanying drawings,

FIG. 1 is an axial cross-section of a thrust bearing formed inaccordance with the present invention,

FIG. 2 is an elevational view partially in section of a machine-toolsuchas a lathe-grinder employing a hearing of the present invention, and

FIG. 3 shows diagrammatically the functioning of the bearing of thepresent invention in three states, namely,

FIG. 4 shows schematically a perspective view of a thrust bearing formedin accordance with the present invention.

DESCRIPTION OF INVENTION For illustration the present invention isdescribed in the form of a thrust or axial slide bearing for thehorizontally rotating shaft of a machine tool such as a universallathe-grinder of conventional type although the application generally tothe support of rotating shafts will be obvious. Such machine is seen inFIG. 2 and in addition to the conventional lathe headstock A andtailstock B, it comprises a housing body 1 set back rearwardly from theaxis thereof, of the stocks A and B. The housing body 1 rests upon a bed13 supported by the machine base l2 and journals a shaft or spindle 2extending parallel to the axis of the headstock and tailstock. A grinderwheel 14 is mounted at the forward and of the spindle 2 while suitabledrive means 15, such as, flywheels, gear box, pulley train connected toa motor, or a motor itself, is connected at the rear end. The specificdetails of such machine are believed to be so well known that furtherdescription here is not believed to be necessary.

The spindle 2 is mounted within the housing body 1 by means ofa pair ofhydrodynamic radial journal bearings 17 located at the forward and rearends. The spindle 2 is preferably necked in one or more coaxial stepsand is provided with the axial or thrust hydrodynamic bearing 16 of thepresent invention, to prevent axial translation. A preferred form ofradial bearings is fully described in the copending application of thesame inventor based upon his Czechoslovak patent appliction No. PV1754-70, filed Mar. 17, 1970.

The-thrust bearing of the present invention is seen in detail in FIG. 1.While this bearing is seen in partial sectional form it will beunderstood that it comprises bodies of revolution about the centrallongitudinal axis of the spindle 2. The bearing comprises a ring shapedannular flange 3 of stainless steel or the like material fixedly mountedfor axial and conjoint rotative movement on the spindle 2. The flange 3is backed by collar 4 hydraulically expanded which is set on the spindle2 and then, after releasing hydraulic expansion there rigidly fixed tolimit the movement of the flange 3 which may also be keyed to thespindle. The flange 3 extends radially from the spindle between a pairof slide rings or discs 5 and 6 which are preferably made of bronze orother similar metal. The discs 5 and 6 are held within an annularsurrounding shell 7 by the provision of mating longitudinally extendingkey and keyways means which permits relative axially movement for atleast the disc 6 but, no rotative movement. Annular washers forming endretaining members 8 and 9 are provided to hold the discs 5 and 6 fromaxially moving external of the shell. The washers 8 and 9 are fastenedto the shell 7 and simultaneously to the machine tool housing body 1 byan elongated bolt fastener 10 which extends longitudinally through allthe members.

The retaining washer 8 holding the front disc 5 (in the direction of thepower source) is rigid and unyielding and it therefore fixes the extentof movement of the disc 5. The rear retaining washer 9 (in the directionof the grinder) however, is resiliently formed with a predeterminedrigidity so as to yield to the movement of its associated slide 6 underapplication of a given pressure. The rear washer 9 is located andcompressed in part between the shell 7 and the body 1 but has its inneredge (i.e. that portion adjacent the disc 6) profiled as a free c'urvedBelleville washer to bear against the slide] The rear washer 9 is madefrom suitable resilient material of predetermined tensile strength andis profiled in a predetermined preselected manner to provide a selectedstatic preloading of a given level of force on the rear face of theslidable disc 6. A degree of rigidity and or biasing pressure suitableand advantageous for the type of use to which the bearing may be putlies within the range of 0.5 to 5K/ micron. Said Belleville washer ismade of usual spring steel and its elasticity is determined such that itmust fit to the formula where P,,,, is axial prestressing of a spindle8,, is the thickness of formed oil film 8, is manufacture axialclearance, and

c is axial rigidity of a retaining washer 9 (rigidity ration between aload and caused deformation) The value (28,, 13,-) is for each bearingknown, the force P must be approximately the same as an outer loadingforce P and it is known too. As consequence of that the value ofrigidity 0 (there rigidity l/elasticity) can be determined exactlybefore assembling.

The inner surfaces of the slidable discs 5 and 6 are substantiallyparallel and each forming a planar functional surface. Each of thesurfaces are formed with lubricating grooves 51 and 61 and cells 11distributing a film of oil between the functional surfaces to providethe fluid film bearing. Each of the surfaces are further provided withbeveled portions forming entering wedges by which the fluid may becompressed. Lubricating oil is supplied from a reservoir (not shown)through suitable channels 71 formed in the shell 7 extending intangential grooves 51 and 61 to the functional surfaces of slide discs 5and 6.

The cells 11 are formed in the face of the discs 5 and 6 by stressing ordeflecting the ring and in that state grinting its surface in planarform. On release of the stress the planar surface takes on a corrugatedconfiguration and the cells 11 are formed. The final corrugated form ofdiscs 5 and 6 surfaces surfaces can be seen to gether with resultingcells 11 in FIG. 4. Other means may, of course, be used.

In operation of the machine tool the spindle 2 is of course, rotatedcausing the oil to be compressed in the entering wedges between theinterfaces of the flange 3 and the discs 5 and 6 respectively. A film isforced between each of the functioning surfaces to form the actualbearing which stresses the slidable disc 6 and washer 9. Hydrodynamicforces are thus created which cause the bearing to axially extend. Sincethe front disc 5 is firmly held in fixed position by the washer 8 thedynamic forces generate a pressure vector acting on the resilient washer9 (actually against its radially curved portion) permitting the movabledisc 6 to slide axially with respect to the spindle. This permits theoil films to vary in thickness and corresponding pressure dependent uponthe forces exerted by the flange 3. The system tends to maintain adynamically balanced state, the result of which is to tend to readjustthe axial position of the spindle so that the flange 3 is provided withan optimum axial clearance from the opposed faces of the discs 5 and 6.The dynamic balance is obtained by proper dimensioning of the resilientcurved (Belleville) portion of the washer 9 by which optimum rigidity isobtained as well as optimum bearing capacity without fear of seizure.

As the speed of the spindle 2 is increased the forces acting on it tendto move it axially and tend to deform the washer 9 still further and indirect response thereto. Thus the disc 6 is shifted still furtheraxially against the resilient washer changing the corresponding pressureand thickness of the films, until a new optimum clearance is obtained.in fact, the increased speed of spindle rotation only acts to morepositively insure the existence of an oil film between the relativelyrotating parts under sufficient pressure to prevent binding or seizingof the parts. It has been found that operation of the spindle can beoptimumly obtained with the range extending from a minimum of 200 RPMsto a maximum of 20,000 RPMs. The provision of the washer 9 which exertsa static biasing or load on the axially movable disc 6 permits and infact, results in automatic readjustment instantaneously on variation ofspindle speed while optimum bearing conditions are maintained.

The functioning of the bearing is depicted in the three stages indicatedin FIG. 3 which schematically indicates by corresponding numerals theoperative elements. In the rest position (i.e., when the spindle is notrotating) as seen in FIG. 3 a the bearing is preadjusted so that aninitial axial clearance is formed between the interfaces of thefunctional surfaces of the disc 6 and the flange 3, when the flange 3 ispreloaded or biased against the fixed disc 5. The spindle is, of course,fully seated in its drive means so that any subsequent axial movementwill be immediately translated against the movable discs 6. The axialclearance is denoted by the symbol 8, and is preferably in the range ofbetween 0 to microns. Little special attention need be made to thestructure of the discs or flange since such a range is easily withinreasonable manufacturing tolerances for such parts. The preadjustment,of course, is obtained by mounting the bearing and spindle initially toeffect the position of the flange 3 flush against the slide disc 5 andseated in its drive means.

As the spindle is rotated, without a load being placed on it, oilbecomes interposed between the functional surfaces (of flange 3 anddiscs 5 and 6 respectively) and a fluid film is formed. The speed of thespindle causes it to axially shift away from the drive means. Thisplaces a new prestressing on the movable discs 6 and the resilientwasher 9 moving them toward the work piece. The spindle moves a distancee while the discs 6 and washer9 are caused to move twice thatdistance-or 2 due to the creation of film 8 between each of theinterfaces of flange 3 and discs 5 and 6. In the absence of externalloading of the spindle the film be tween the flange 3 and each of thediscs 5 and 6 is equal in thickness and in pressure. The thickness beingdenoted by 8 the pressure by P The generation of this balanced clearanceis enabled by the resilient deformation of the retaining washer 9 whichpermits the flange 3 and movable disc 6 to move axially in response topressure P hydrodynamically produced by the compression and movement ofthe fluid (oil) molecules between the functional surfaces. Thehydrodynamic force developed during the rotation of the spindle withoutexternalload, as seen in FIG. 3 b, can be calculated by the followingequation:

where:

25,, 5. is the distance the resilient washer 9 will flex, and

C denotes the rigidity of the resilient washer 9.

(As previously noted preferably somewhere in the range of 0.5 to 5 K/micron.)

The flexing of the washer 9 results in the increase of the springloading or biasing of the washer 9 on the disc 6 equal to a force P (thedeformation force of the washer 9) which to balance the system equalsthe force P On increasing the speed and applying an operating load onthe spindle 2 (i.e., by placing the grinding wheel 14 into work positionon a work piece) an external load or force P is impressed on thebearing. Said force P can act in both axial direction, wherein thedirection shown in FIG. 3 is most disadvantageous. The spindle isthereafter shifted into a new balanced position responsive to theexternal impressed force which consequence automatic readjustment of thespindle is creates a new hydrodynamic ratio and balance as seen in FIG.3 c. The thickness of the original lubricating oil film is furtherchanged by a value Z responsive to the direction of movement ordeflection of the spindle 2. In the case shown in FIG. 3 c the spindlemoves from right to left under the external load establishing nonuniformfilms 8, and 6,, respectively. As a result of the shift and the changein thickness of the oil film different pressures P and P are developedbetween the flange 3 and each of the faces of the discs 6 and 5respectively. The hydrodynamic forces represented by the pressurevectors P and P must be in balance with the outer load P as well as withthe deformation force P acting from the resilient washer. Consequentlythe balanced state of the axial hydrodynamic thrust bearing is expressedas the equation PT=PDI PDH however,

(P (P being a force with which a washer 9 acts on a slide disc 6 it isalso an axial prestressing of a bear-' ing P P but since the deformationof the resil ient member under external load condition is the sum ofthe'deformation force vector under no external load (P or P plus thechange in oil film thickness Z under external load multiplied by therigidity factor C of the washer 9, thus Under external load conditionstherefore, the spindle shifts in response to the resultant externalforce vector maintaining a film of oil between the interfaces of flange3 and the slide discs 5 and 6 under a bearing pressure responsivethereto, to maintain a balanced condition. The equations mathematicallydescribe the movement or vector of the forces within the bearing of thepresent invention. They are basic for the calcula' tion of optimumparameters of the bearing structure in order to obtain optimum function.The rigidity of the resilient washer 9, bearing capacity and precisionof the parts can all be derived through these formulae.

During continued operation of the grinder-lathe the spindle 2 is subjectto varying external load conditions. The varying load conditions are afunction of the speed of rotation of the spindle and the external load.As a continually obtained as a response to all the varying load factors.

It will thus be seen that optimum bearing conditions are continually metby the present invention permitting the operation of the spindle at anyspeed and even under varying speed conditions. The bearing is providedwith a constantly variyble elasticity responsive to the load whichpermits both low speed and extremely high speed operation. Since abearing film exists at all times the thickness of which is responsive tothe speed, the functional forces between rotating surfaces as well asthe temperatures generated are minimal. Thus long operation even undervarying loads can be made with minimum danger of damage or seizure. Ithas been found that sudden variations in a range ratio of minimum tomaximum revolution of 1:100 can be easily tolerated with effect on thebearing.

It will be also seen that the relatively moving parts need not be madewith great precision and ordinary tolerance are permitted. Thus thedevice is not only simple and effective but also inexpensive.

Various modifications and changes may be made. Equivalent structuralforms and materials may be used. In addition to those given herein,those skilled in the art will recognize other modifications and changes.The disclosure is to be taken as illustrative only, and not limiting ofthe present invention.

What is claimed:

1. A hydrodynamic thrust bearing for a rotating spin dle and the likecomprising a cylindrical shell adapted to surround and be spaced fromsaid spindle, and annular flange adapted to be secured to said spindlefor conjoint movement therewith and extending towards said shell, afirst annular disc, adapted to surround said spindle located on one sideof said flange and fixedly supported by said shell, a second, annulardisc adapted to surround said spindle and located on the other side ofsaid flange said second disc being slidably supported by said shell forfree axial movement relative to said spindle during rotation thereof,means for resiliently biasing said movable disc normally toward saidannular flange, and means for supplying bearing fluid between saidflange and said discs respectively, said movable disc providinghydrodynamic pressure within said shell during rotation of said spindle.

2. The bearing according to claim 1 wherein the means for supportingsaid discs by said shell include a washer associated with each of saiddiscs, and fastening means for fixedly mounting said washers to the endsof said shell to retain said discs against axial movement external ofsaid shell.

3. The bearing according to claim 2 wherein said washer associated withsaid movable disc is at least, in part, resiliently formed and bearsforcibly on said disc.

4. The bearing according to claim 3 wherein said resilient washercomprises a Belleville spring.

5. The bearing according to claim 1 wherein said shell and said discsare provided with mating keys and keyways extending along a longitudinalaxis thereby securing said discs against rotation.

6. The bearing according to claim 1 wherein the faces of said discs areformed with grooves for distributing bearing fluid and a beveled portionforming an entering wedge between said discs and said flange for saidfluid.

7. The bearing according to claim 1 including means for securing saidshell against movement to thereby provide a fixed base for said bearing.

8. The bearing according to claim 1 wherein the ratio of axial loadingto the caused axial deflection of said resilient biasing means is withinthe range of 0.5 to 5 K /micron.

9. The bearing according to claim 1 wherein said movable disc is locatedon the side of said disc corresponding to the end of said spindle onwhich external work load is applied.

10. A fluid bearing system for absorbing the thrusting load ofarotatable shaft or the like comprising an annular flange, adapted to besecured to said shaft and located between a first fixed annular disc anda second freely axially movable annular disc each of said discs beingadapted to surround said spindle, means axially prestressing saidmovable disc toward said flange and means supplying fluid to the facesbetween said flange and disc respectively, to provide a hydrodynamicaction within said bearing whereby on rotation of said spindle saiddiscs cause displacement of said fluid and an increase of pressurebetween said faces to cause said movable disc to distend axially.

11. The system according to claim 10 wherein said movable disc isprestressed under rotating conditions of said spindle with no externalload according to the formula m; a i) CM where P is the prestressingforce (hydrodynamical) 8, is the clearance between the flange and thediscs under vector static conditions 5,, is the clearance under rotatingconditions and C is the rate constant for the rigidity or spring rate ofthe biasing or prestressing member.

12. The system according to claim 10 wherein under external loadconditions the bearing is maintained at hydrodynamic balance accordingto the following formula where P is the resultant force of the externalload, P is the force between the face of the flange and the slidabledisc, P is the force between the face of the flange and the fixed disc,and

where P is the total stressing of the movable member, and

Z is the displacement of a slidable disc with rotation under the effectof an outer load.

1. A hydrodynamic thrust bearing for a rotating spindle and the likecomprising a cylindrical shell adapted to surround and be spaced fromsaid spindle, and annular flange adapted to be secured to said spindlefor conjoint movement therewith and extending towards said shell, afirst annular disc, adapted to surround said spindle located on one sideof said flange and fixedly supported by said shell, a second, annulardisc adapted to surround said spindle and located on the other side ofsaid flange said second disc being slidably supported by said shell forfree axial movement relative to said spindle during rotation thereof,means for resiliently biasing said movable disc normally toward saidannular flange, and means for supplying bearing fluid between saidflange and said discs respectively, said movable disc providinghydrodynamic pressure within said shell during rotation of said spindle.2. The bearing according to claim 1 wherein the means for supportingsaid discs by said shell include a washer associated with each of saiddiscs, and fastening means for fixedly mounting said washers to the endsof said shell to retain said discs against axial movement external ofsaid shell.
 3. The bearing according to claim 2 wherein said washerassociated with said movable disc is at least, in part, resilientlyformed and bears forcibly on said disc.
 4. The bearing according toclaim 3 wherein said resiliEnt washer comprises a Belleville spring. 5.The bearing according to claim 1 wherein said shell and said discs areprovided with mating keys and keyways extending along a longitudinalaxis thereby securing said discs against rotation.
 6. The bearingaccording to claim 1 wherein the faces of said discs are formed withgrooves for distributing bearing fluid and a beveled portion forming anentering wedge between said discs and said flange for said fluid.
 7. Thebearing according to claim 1 including means for securing said shellagainst movement to thereby provide a fixed base for said bearing. 8.The bearing according to claim 1 wherein the ratio of axial loading tothe caused axial deflection of said resilient biasing means is withinthe range of 0.5 to 5 Kp/micron.
 9. The bearing according to claim 1wherein said movable disc is located on the side of said disccorresponding to the end of said spindle on which external work load isapplied.
 10. A fluid bearing system for absorbing the thrusting load ofa rotatable shaft or the like comprising an annular flange, adapted tobe secured to said shaft and located between a first fixed annular discand a second freely axially movable annular disc each of said discsbeing adapted to surround said spindle, means axially prestressing saidmovable disc toward said flange and means supplying fluid to the facesbetween said flange and disc respectively, to provide a hydrodynamicaction within said bearing whereby on rotation of said spindle saiddiscs cause displacement of said fluid and an increase of pressurebetween said faces to cause said movable disc to distend axially. 11.The system according to claim 10 wherein said movable disc isprestressed under rotating conditions of said spindle with no externalload according to the formula PDo (2 delta o - delta i) . CM where PDois the prestressing force (hydrodynamical) delta i is the clearancebetween the flange and the discs under vector static conditions delta ois the clearance under rotating conditions and CM is the rate constantfor the rigidity or spring rate of the biasing or prestressing member.12. The system according to claim 10 wherein under external loadconditions the bearing is maintained at hydrodynamic balance accordingto the following formula PT PD1 - PD11 where PT is the resultant forceof the external load, PD1 is the force between the face of the flangeand the slidable disc, PD11 is the force between the face of the flangeand the fixed disc, and PD1 PM PDo + Z . CM where PM is the totalstressing of the movable member, and Z is the displacement of a slidabledisc with rotation under the effect of an outer load.